The present invention generally relates to the control of fuel vapour generated within an internal combustion engine installation, and in particular to a method for determining the amount of fuel vapour purged from a fuel vapour collection device of the engine installation.
The current emission regulations in many countries require the evaporative emissions from the fuel supply system of the internal combustion engines of motor vehicles to be controlled to thereby eliminate or substantially reduce the amount of fuel released into the atmosphere by such vapours. Accordingly, it is normal practice to fit a fuel vapour collection device to the vehicle to adsorb evaporative emissions from the fuel supply system under all conditions that the vehicle experiences. This fuel vapour collection device is usually of the activated carbon type and is commonly referred to as the xe2x80x9ccarbon canisterxe2x80x9d. Such a fuel vapour collection device operates on the principle of physical adsorption of fuel vapour onto the activated carbon.
The fuel vapour collection device generally has a limited capacity for storing fuel vapour and must therefore be purged to some extent of its contents in the course of vehicle operation. The accumulated fuel vapour is normally purged into the intake manifold of the engine by way of air drawn through the fuel vapour collection device, the purged fuel vapour being subsequently combusted within the engine. The amount of fuel vapour being purged from the fuel vapour collection device can however vary significantly for any given purge air flow rate generally depending on the saturation level in the fuel vapour collection device. As the amount of purged fuel vapour is typically not able to be measured in systems not having an air/fuel ratio feedback mechanism (commonly known as open loop systems), the engine control system for such open loop systems generally cannot compensate for the increased fuelling rate to the engine. This can cause an increase in the engine torque which may result in a higher engine speed at idle or an increase in the vehicle speed off idle. Under severe conditions, the engine operation can become unstable because the actual air fuel ratio within the engine cylinders is markedly different from the air fuel ratio mapped by the engine control system.
In the Applicant""s U.S. Pat. No. 5,245,974 there is described a fuel vapour control system for an internal combustion engine, the details of which are incorporated herein by reference. This document discloses an internal combustion engine installation having a fuel vapour collection device for removing the fuel vapour from the evaporative emissions generated within the fuel supply system. The engine includes a dual fluid fuel injection system with an air compressor supplying compressed air to the fuel injection system. The fuel vapour collection device is periodically purged of accumulated fuel vapour by way of drawing air through the fuel vapour collection device using the air compressor. The air compressor then supplies the air which now carries the fuel vapour to the fuel injection system where the air is subsequently injected into the combustion chambers of the engine resulting in combustion of the purged fuel vapour. Although the stratification within the cylinder will remain largely unaltered by the addition of the purged fuel through the injector, this patent does not particularly address the problem of lack of knowledge of the amount of fuel being supplied from the fuel vapour collection device.
A proposal for dealing with this problem is described in the Applicant""s International Publication No. WO 00/01663, the details of which are also incorporated herein by reference. This document describes a method of controlling the flow rate of a purge flow passing through a fuel vapour collection device by controlling the opening of a flow control valve located between the vapour collection device and the engine. The method controls the flow control valve as a function of engine operating conditions. However, the described method does not actually determine the amount of fuel vapour in the purge flow to the engine. An iterative method for providing an estimation of the fuel flow rate based on empirical data is actually used. This application also describes a method of determining the amount of fuel vapour being purged during closed loop operation of the engine. The engine is typically operated in this manner when the engine is at idle. It may also be possible to operate the engine under closed loop control when operating at stoichiometric air/fuel ratio conditions. However, at other engine loads such as at partial loads, it is necessary to operate the engine under open loop control where the fuel purge flow rate cannot be directly determined.
It would therefore be advantageous to be able to determine the actual amount of fuel within the purge flow to the engine under most, if not all engine operating conditions.
With this in mind, an object of the present invention is to provide an improved method for determining a purge fuel mass flow rate from a fuel vapour control system to an internal combustion engine, at least under most engine operating conditions.
According to the present invention, there is provided a method for determining a purge fuel mass flow rate from a fuel vapour control system to an internal combustion engine having a compressor for delivering purge gas from the fuel vapour control system to the engine, the method including:
determining the temperature rise of the purge gas passing through the compressor;
determining the specific heat ratio of the purge gas as a function of the temperature rise; and
determining the purge fuel mass flow rate as a function of the specific heat ratio of the purge gas.
Preferably, the fuel vapour control system includes an air/fuel separation means for collecting fuel vapour generated within the engine. Preferably, the compressor is arranged to deliver purge gas from the air/fuel separation means to the engine. It is however contemplated that the compressor may deliver to the engine purge gas or fuel vapour generated or present anywhere within the engine.
As the purge fuel mass flow rate is determined as a function of the temperature rise of the purge gas as it passes though the compressor, the determination of the purge fuel mass flow rate is independent of the engine operating conditions. Therefore, the purge fuel mass flow rate can be determined under most, if not all engine loads and speeds.
The specific heat ratio of the purge gas varies in dependence on the purge fuel concentration of the purge gas. Further, the specific heat ratio of purge fuel is significantly different from the specific heat ratio of air. For example, the specific heat ratio of air is about 1.4, whereas typical purge fuel species such as C3H8, C4H10, and C5H14 have specific heat ratios of between 1.06 to 1.11. Generally, the higher the molecular weight of the gas, the lower the value of the specific heat ratio.
Therefore, as the concentration of purge fuel or fuel vapour within the purge gas increases, the specific heat ratio of the purge gas decreases. Hence, by monitoring the change in the adiabatic temperature rise of the gas passing through the compressor, the purge fuel mass flow rate may be determined. That is, as the compressor goes from delivering solely air to the engine to delivering both air and purge fuel, the specific heat ratio of the gas passing through the compressor will change.
Most positive displacement compressors provide approximately adiabatic compression of gas passing through the compressor. However, the compressor can never provide truly adiabatic compression under actual operating conditions due to heat losses within the compressor and in general a real compressor provides compression which is neither adiabatic nor isothermal. It can be modelled however by using the polytropic compression equation:
TOUT=TINxc3x97PRnl(nxe2x88x92l)
where
TOUT is the compressor discharge temperature;
TIN is the compressor inlet temperature;
PR is the pressure ratio across the compressor; and
n is the polytropic index.
The polytropic index for positive displacement compressors using air as the fluid being delivered is typically 1.3. If the compressor was ideal and provided adiabatic compression, then n=Cp/Cv (the specific heat ratio of the purge gas) which equals 1.4 for air. The difference reflects the fact that the compressor is not perfect and does have losses.
For a compressor which has a gas of variable composition passing through it, the above equation may be modified as follows:       T    OUT    =            T      IN        xc3x97          PR              k        ⁢                  (                      γ                          γ              -              1                                )                    
where
xcex3 is the specific heat ratio (Cp/Cv) for the mixture.
The value k reflects the fact that it is a real process and therefore allows for losses to occur. The value of k may be determined for a particular compressor. Hence the specific heat ratio of the purge gas passing through the compressor may be determined by the following equation:       T    OUT    =            T      IN        xc3x97                  PR        k            ⁢              (                                            C              p                        /                          C              v                                            (                                                            C                  p                                /                                  C                  v                                            -              1                        )                          )            
As the pressure ratio across the compressor is generally constant, the specific heat ratio of the purge gas can be determined by measuring the compressor discharge temperature. This temperature can for example be measured by a temperature sensor such as a thermistor located downstream of the discharge port of the compressor. In the Applicant""s engine installations, the air inlet temperature is generally measured at the air inlet or intake manifold for engine control purposes. The compressor inlet temperature can essentially be set as being the same as the temperature of the air at the intake manifold, with any temperature increase of the air between the intake manifold and the compressor inlet being likely to be minimal. It is however also envisaged that a further temperature sensor could be provided immediately upstream of the compressor intake for greater accuracy.
An average discharge temperature may be determined by electronically filtering the signal from the temperature sensor. Alternatively, the temperature sensor may be placed a sufficient distance away from the compressor discharge port. It is also preferable for a temperature sensor having a relatively low dynamic response to be used to thereby avoid the need to further dampen the signal from the sensor.
As mentioned above, the compressor does not produce a truly adiabatic compression process due to heat losses to the compressor components and housing surrounding the compression chamber of the compressor. For example, in the case of a piston compressor, heat losses may occur to the cylinder wall, the compressor head and the compressor piston. The method according to the present invention may therefore include compensation means to account for the effect of the abovenoted heat losses by monitoring the coolant temperature of the engine and adding a compensation factor to the abovenoted determination. For example, the difference in the coolant temperature between the nominal value at which a calibration was done and the actual value at current engine operating conditions could be measured and simply added as an offset.
The heat loss per compressor cycle will typically also be inversely proportional to the operational speed of the compressor. To this end, the method may further include mapping or calculating the heat loss at different compressor speeds. For example, a series of tests could be done on a compressor installation from which an experienced relationship could be developed to reflect the variation with compressor speed. It should also be noted that the compressor speed has a direct effect on the temperature rise across the compressor and therefore a suitable compensation factor may also be required to take into account changes in the compressor speed. Such a compensation factor could either be mapped or have a suitable algorithm determined therefore. In essence the simplest approach would be to have a look-up map which provides compensation for compressor speed as this would automatically allow for the heat loss per compressor cycle varying with compressor speed.
As noted above, the inlet temperature to the compressor intake can be assumed to be equal to the temperature of the air at the engine intake. If the inlet temperature variations are likely to be significant, then a non-linear compensation for the intake temperature could be provided in the form of a further calculation or by the inclusion of a compensation look-up map provided in an ECU of the engine.
The compressor temperature characteristics may alter over time because of degradation of the compressor performance due to, for example, ring wear, valve leakage and so on. This may progressively decrease the accuracy of the purge fuel mass flow rate determination. Furthermore, the pressure ratio will change if there is a restriction to the compressor intake gas flow. This can arise due to, for example, a dirty air filter.
When the engine is operating under closed loop fuelling control (ie: typically at idle or at stoichometric combustion conditions), the purge fuel per cycle to the engine can be measured by alternative means. For example, such a system is disclosed in the Applicant""s U.S. Pat. No. 5,806,304, the contents of which are included herein by reference. To this end, the method of the present invention may include comparing the purge fuel mass flow rate determined during closed loop fuelling control with the purge fuel mass flow rate determined by the method according to the present invention. This enables the determination method to be checked for accuracy and adjusted or adapted as required. For example, such an approach could be used to allow for compensation for slightly different fuel vapour compositions due to different grades of fuel (eg. ULP vs PULP vs Super, different RVP""s), variations from refineries (typically quite small), and variations in conditions which produce the vapour. The actual change in heat capacity ratios arising from such variations is not expected to be very significant (eg: around 5% maximum), but in the interests of greater accuracy, these could be taken into account if desired. Such a comparison routine could also for example, counter the effect produced by a restriction in the intake of the compressor or by mechanical degradation of the compressor.
Under certain circumstances or in relation to particular engine applications, the method according to the present invention may be used together with other known purge fuel mass flow rate determination means to calculate the purge fuel mass flow rate across all regions of engine operation. For example, under closed loop fuelling control operation, either of the methods alluded to hereinbefore may be used to calculate purge fuel mass flow rate, whilst for operation say during partial load (ie: generally open loop fuelling control), the method according to the present invention could be used to determine purge fuel mass flow rate.
The method according to the present invention has significant practical advantages over known purge fuel mass flow rate control methods. In particular, with respect to the Applicant""s International Publication No. WO 00/01663, the method according to the present invention can eliminate the need for the flow control valve and the associated system used to control and drive the control valve. This can result in significant cost savings. The method of the present invention is a comparatively low cost system in that it mainly relies on a low cost thermistor at the compressor outlet as the primary additional hardware. Furthermore, because the purge fuel mass flow rate is being continuously determined, there is no need to attempt to predict this mass flow rate as in the method described in the abovenoted international application. That is, the purge fuel mass flow rate determination method according to the present invention essentially provides for closed loop fuel vapour purge operation for all engine operating conditions.
The method according to the present invention is particularly applicable for four stroke engines having a fuel vapour control system including an air/fuel separation means and a compressor for delivering purged gas from that separation means to the engine. It is however also possible for the method to be used on a two stroke engine having a similar fuel vapour control system.
Conveniently, the compressor forms part of a dual fluid fuel injection system for the engine wherein metered quantities of fuel are delivered to the engine entrained in gas, typically air, supplied by the compressor. Such a dual fluid fuel injection system is for example disclosed in the Applicant""s U.S. Pat. No. 4,934,329, the contents of which are included herein by reference. Conveniently, the fuel injection system is configured such that the fuel entrained in air is delivered directly into the combustion chamber(s) of the engine. Hence, any purge fuel or gas delivered to the engine by way of the compressor will be delivered directly into the combustion chamber(s) of the engine.
Conveniently, where the fuel injection system may include air pressure regulation means for dumping excess air that is delivered by the compressor, the method may be sophisticated enough to compensate for any fuel vapour which is recirculated back through the compressor. For example, an air pressure regulator may be configured to dump excess air delivered by the compressor under certain running conditions back to the engine air intake or the intake of the compressor. Hence, if a quantity of fuel vapour was to be purged through the compressor during such running conditions, some of the fuel vapour would be delivered to the engine together with the air delivered by the compressor whilst some of the fuel vapour would invariably be recirculated with the excess air returned by the air pressure regulator. However, as it would typically be possible to determine the volume of air delivered by the fuel injection system delivery injectors to the engine, it would equally be possible to determine the volume of the air and fuel vapour mix actually delivered to the engine. Therefore the volume of air and fuel vapour regulated by the air pressure regulator back into the engine air intake or compressor intake could be determined. The method according to the present invention could hence be arranged to compensate on the basis of any such recirculated fuel vapour (eg. air pressure regulation compensation factor) so as to maintain the accuracy of the determination of purge fuel mass flow rate.
In certain applications, the fuel injection system may be configured to allow for throttling of the intake of the air compressor to improve the efficiency of the overall system. However, throttling of the compressor intake also has the effect of changing the pressure ratio across the compressor (ie. PR may not necessarily be constant). Conveniently, the system is able to compensate for such variable PR values such that an accurate determination of the purge fuel mass flow rate may be made. For example, a suitable compensation factor could be determined by a method of calculation or a method of measurement.
To enable compensation by way of measurement, a suitable pressure sensor may be provided at the compressor intake to determine the throttled air pressure. That is, a measurement of the air pressure downstream of a particular throttling means which is used (eg. a suitable butterfly valve in the compressor intake passage) is made. A further air pressure sensor may then be provided downstream of an outlet of the compressor and the readings from each sensor compared in order to determine the pressure ratio across the compressor due to a particular degree of throttling. Alternatively, rather then provide a second air pressure sensor downstream of the compressor outlet, if the air pressure downstream of the compressor is being regulated to a predetermined value (eg. one necessary for satisfactory operation of the fuel injection system), the reading from the first sensor may be compared to this fixed value in order to determine the pressure ratio across the compressor for the particular degree of throttling applied.
To enable compensation via a method of calculation, the relationship between the degree of throttling of the compressor intake and the pressure ratio across the compressor could simply be mapped at the point at which the. engine is initially calibrated. This information can then be used to create an appropriate lookup table such that during operation of the engine, an engine control system is able to use the lookup table to determine a PR value corresponding to any degree of throttling which is applied via the throttling means
Hence, by way of either of these methods, a suitable PR value can be calculated for a certain degree of throttling of the compressor intake and this PR value can be used to enable accurate determination of the purge fuel mass flow rate.